Internal Combustion Engine Parts n 1125
2. Principal Parts of an I. C.
3. Cylinder and Cylinder
4. Design of a Cylinder.
6. Design Considerations for
7. Material for Pistons.
8. Piston Head or Crown .
9. Piston Rings.
10. Piston Barrel.
11. Piston skirt.
12. Piston Pin.
13. Connecting Rod.
14. Forces Acting on the
Connecting Rod. 32.1 Introduction
15. Design of Connecting Rod.
16. Crankshaft. As the name implies, the internal combustion engines
17. Material and Manufacture (briefly written as I. C. engines) are those engines in which
of Crankshafts. the combustion of fuel takes place inside the engine cylinder.
18. Bearing Pressures and The I.C. engines use either petrol or diesel as their fuel. In
Stresses in Crankshafts. petrol engines (also called spark ignition engines or S.I
19. Design Procedure for engines), the correct proportion of air and petrol is mixed
Crankshaft. in the carburettor and fed to engine cylinder where it is
20. Design for Centre ignited by means of a spark produced at the spark plug. In
Crankshaft. diesel engines (also called compression ignition engines
21. Side or Overhung or C.I engines), only air is supplied to the engine cylinder
during suction stroke and it is compressed to a very high
22. Valve Gear Mechanism.
23. Valves. pressure, thereby raising its temperature from 600°C to
24. Rocker Arm. 1000°C. The desired quantity of fuel (diesel) is now injected
into the engine cylinder in the form of a very fine spray and
gets ignited when comes in contact with the hot air.
The operating cycle of an I.C. engine may be
completed either by the two strokes or four strokes of the
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piston. Thus, an engine which requires two strokes of the piston or one complete revolution of the
crankshaft to complete the cycle, is known as two stroke engine. An engine which requires four
strokes of the piston or two complete revolutions of the crankshaft to complete the cycle, is known as
four stroke engine.
The two stroke petrol engines are generally employed in very light vehicles such as scooters,
motor cycles and three wheelers. The two stroke diesel engines are generally employed in marine
The four stroke petrol engines are generally employed in light vehicles such as cars, jeeps and
also in aeroplanes. The four stroke diesel engines are generally employed in heavy duty vehicles such
as buses, trucks, tractors, diesel locomotive and in the earth moving machinery.
32.2 Principal Parts of an Engine
The principal parts of an I.C engine, as shown in Fig. 32.1 are as follows :
1. Cylinder and cylinder liner, 2. Piston, piston rings and piston pin or gudgeon pin, 3. Connecting
rod with small and big end bearing, 4. Crank, crankshaft and crank pin, and 5. Valve gear mechanism.
The design of the above mentioned principal parts are discussed, in detail, in the following
Fig. 32.1. Internal combustion engine parts.
32.3 Cylinder and Cylinder Liner
The function of a cylinder is to retain the working fluid and to guide the piston. The cylinders
are usually made of cast iron or cast steel. Since the cylinder has to withstand high temperature due to
the combustion of fuel, therefore, some arrangement must be provided to cool the cylinder. The
single cylinder engines (such as scooters and motorcycles) are generally air cooled. They are provided
with fins around the cylinder. The multi-cylinder engines (such as of cars) are provided with water
jackets around the cylinders to cool it. In smaller engines. the cylinder, water jacket and the frame are
- Internal Combustion Engine Parts n 1127
made as one piece, but for all the larger engines, these parts are manufactured separately. The cylinders
are provided with cylinder liners so that in case of wear, they can be easily replaced. The cylinder
liners are of the following two types :
1. Dry liner, and 2. Wet liner.
Fig. 32.2. Dry and wet liner.
A cylinder liner which does not have any direct contact with the engine cooling water, is known
as dry liner, as shown in Fig. 32.2 (a). A cylinder liner which have its outer surface in direct contact
with the engine cooling water, is known as wet liner, as shown in Fig. 32.2 (b).
The cylinder liners are made from good quality close grained cast iron (i.e. pearlitic cast iron),
nickel cast iron, nickel chromium cast iron. In some cases, nickel chromium cast steel with molybdenum
may be used. The inner surface of the liner should be properly heat-treated in order to obtain a hard
surface to reduce wear.
32.4 Design of a Cylinder
In designing a cylinder for an I. C. engine, it is required to determine the following values :
1. Thickness of the cylinder wall. The cylinder wall is subjected to gas pressure and the piston
side thrust. The gas pressure produces the following two types of stresses :
(a) Longitudinal stress, and (b) Circumferential stress.
Piston ring seals the piston to
prevent gases escaping
Camshaft controls the
Valve lets fuel and air in and
exhaust gases out
Belt drives alternator to supply
electricity to spark plugs.
Crankshaft turns the Dip stick to check oil level
piston action into rotation Sump is filled with oil to
Oil is pumped up into
Piston cylinders to lubricate pistons
The above picture shows crankshaft, pistons and cylinder of a 4-stroke petrol engine.
- 1128 n A Textbook of Machine Design
Since these two stressess act at right angles to each other, therefore, the net stress in each
direction is reduced.
The piston side thrust tends to bend the cylinder wall, but the stress in the wall due to side thrust
is very small and hence it may be neglected.
Let D0 = Outside diameter of the cylinder in mm,
D = Inside diameter of the cylinder in mm,
p = Maximum pressure inside the engine cylinder in N/mm2,
t = Thickness of the cylinder wall in mm, and
1/m = Poisson’s ratio. It is usually taken as 0.25.
The apparent longitudinal stress is given by
Force × D2 × p D2 . p
σl = = 4 =
[( D0 ) 2 − D 2 ] ( D0 ) − D
Area 2 2
and the apparent circumferential stresss is given by
Force D × l × p D × p
σc = = =
Area 2t × l 2t
... (where l is the length of the cylinder and area is the projected area)
∴ Net longitudinal stress = σl −
and net circumferential stress = σc −
The thickness of a cylinder wall (t) is usually obtained by using a thin cylindrical formula,i.e.,
p × D
t = 2σ + C
where p = Maximum pressure inside the cylinder in N/mm2,
D = Inside diameter of the cylinder or cylinder bore in mm,
σc = Permissible circumferential or hoop stress for the cylinder material
in MPa or N/mm2. Its value may be taken from 35 MPa to
100 MPa depending upon the size and material of the cylinder.
C = Allowance for reboring.
The allowance for reboring (C ) depending upon the cylinder bore (D) for I. C. engines is given
in the following table :
Table 32.1. Allowance for reboring for I. C. engine cylinders.
D (mm) 75 100 150 200 250 300 350 400 450 500
C (mm) 1.5 2.4 4.0 6.3 8.0 9.5 11.0 12.5 12.5 12.5
The thickness of the cylinder wall usually varies from 4.5 mm to 25 mm or more depending
upon the size of the cylinder. The thickness of the cylinder wall (t) may also be obtained from the
following empirical relation, i.e.
t = 0.045 D + 1.6 mm
The other empirical relations are as follows :
Thickness of the dry liner
= 0.03 D to 0.035 D
- Internal Combustion Engine Parts n 1129
Thickness of the water jacket wall
= 0.032 D + 1.6 mm or t / 3 m for bigger cylinders and 3t /4 for
Water space between the outer cylinder wall and inner jacket wall
= 10 mm for a 75 mm cylinder to 75 mm for a 750 mm cylinder
or 0.08 D + 6.5 mm
2. Bore and length of the cylinder. The bore (i.e. inner diameter) and length of the cylinder may
be determined as discussed below :
Let pm = Indicated mean effective pressure in N/mm2,
D = Cylinder bore in mm,
A = Cross-sectional area of the cylinder in mm2,
= π D2/4
l = Length of stroke in metres,
N = Speed of the engine in r.p.m., and
n = Number of working strokes per min
= N, for two stroke engine
= N/2, for four stroke engine.
We know that the power produced inside the engine cylinder, i.e. indicated power,
p × l ×A × n
I .P. = m watts
From this expression, the bore (D) and length of stroke (l) is determined. The length of stroke is
generally taken as 1.25 D to 2D.
Since there is a clearance on both sides of the cylinder, therefore length of the cylinder is taken
as 15 percent greater than the length of stroke. In other words,
Length of the cylinder, L = 1.15 × Length of stroke = 1.15 l
Notes : (a) If the power developed at the crankshaft, i.e. brake power (B. P.) and the mechanical efficiency (ηm)
of the engine is known, then
I.P. = η
(b) The maximum gas pressure ( p ) may be taken as 9 to 10 times the mean effective pressure ( pm).
3. Cylinder flange and studs. The cylinders are cast integral with the upper half of the crank-
case or they are attached to the crankcase by means of a flange with studs or bolts and nuts. The
cylinder flange is integral with the cylinder and should be made thicker than the cylinder wall. The
flange thickness should be taken as 1.2 t to 1.4 t, where t is the thickness of cylinder wall.
The diameter of the studs or bolts may be obtained by equating the gas load due to the maximum
pressure in the cylinder to the resisting force offered by all the studs or bolts. Mathematically,
× D2 . p = ns × ( d ) 2 σt
4 4 c
where D = Cylinder bore in mm,
p = Maximum pressure in N/mm2,
ns = Number of studs. It may be taken as 0.01 D + 4 to 0.02 D + 4
dc = Core or minor diameter, i.e. diameter at the root of the thread in
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σt = Allowable tensile stress for the material of studs or bolts in MPa or
N/mm2. It may be taken as 35 to 70 MPa.
The nominal or major diameter of the stud or bolt (d ) usually lies between 0.75 tf to tf, where
tf is the thickness of flange. In no case, a stud or bolt less than 16 mm diameter should be used.
The distance of the flange from the centre of the hole for the stud or bolt should not be less than
d + 6 mm and not more than 1.5 d, where d is the nominal diameter of the stud or bolt.
In order to make a leak proof joint, the pitch of the studs or bolts should lie between 19 d to
28.5 d , where d is in mm.
4. Cylinder head. Usually, a separate cylinder head or cover is provided with most of the engines.
It is, usually, made of box type section of considerable depth to accommodate ports for air and gas
passages, inlet valve, exhaust valve and spark plug (in case of petrol engines) or atomiser at the centre
of the cover (in case of diesel engines).
The cylinder head may be approximately taken as a flat circular plate whose thickness (th) may
be determined from the following relation :
th = D
where D = Cylinder bore in mm,
p = Maximum pressure inside the cylinder in N/mm2,
σc = Allowable circumferential stress in MPa or N/mm2. It may be taken
as 30 to 50 MPa, and
C = Constant whose value is taken as 0.1.
The studs or bolts are screwed up tightly alongwith a metal gasket or asbestos packing to provide
a leak proof joint between the cylinder and cylinder head. The tightness of the joint also depends
upon the pitch of the bolts or studs, which should lie between 19 d to 28.5 d . The pitch circle
diameter (Dp) is usually taken as D + 3d. The studs or bolts are designed in the same way as discussed
Example 32.1. A four stroke diesel engine has the following specifications :
Brake power = 5 kW ; Speed = 1200 r.p.m. ; Indicated mean effective pressure = 0.35 N / mm 2 ;
Mechanical efficiency = 80 %.
Determine : 1. bore and length of the cylinder ; 2. thickness of the cylinder head ; and 3. size of
studs for the cylinder head.
Ignition 4-Stroke Petrol Engine
Inlet Spark plug Hot gases expand and force Terminal
valve Piston causes a the piston down
spark Exhaust Ceramic
Crankshaft valve insulator
1. Intake 2. Compression 3. Power 4. Exhaust, 5. Spark plug
- Internal Combustion Engine Parts n 1131
Solution. Given: B.P. = 5kW = 5000 W ; N = 1200 r.p.m. or n = N / 2 = 600 ;
pm = 0.35 N/mm2; ηm = 80% = 0.8
1. Bore and length of cylinder
Let D = Bore of the cylinder in mm,
A = Cross-sectional area of the cylinder = × D2 mm2
l = Length of the stroke in m.
= 1.5 D mm = 1.5 D / 1000 m ....(Assume)
We know that the indicated power,
I.P = B.P. / ηm = 5000 / 0.8 = 6250 W
We also know that the indicated power (I.P.),
pm . l . A . n 0.35 × 1.5D × π D 2 × 600
6250 = = = 4.12 × 10–3 D3
60 60 × 1000 × 4
...( ∵ For four stroke engine, n = N/2)
∴ D3 = 6250 / 4.12 × 10–3 = 1517 × 103 or D = 115 mm Ans.
and l = 1.5 D = 1.5 × 115 = 172.5 mm
Taking a clearance on both sides of the cylinder equal to 15% of the stroke, therefore length of
L = 1.15 l = 1.15 × 172.5 = 198 say 200 mm Ans.
2. Thickness of the cylinder head
Since the maximum pressure ( p) in the engine cylinder is taken as 9 to 10 times the mean
effective pressure ( pm), therefore let us take
p = 9 pm = 9 × 0.35 = 3.15 N/mm2
We know that thickness of the cyclinder head,
C.p 0.1 × 3.15
th = D σ = 115 42 = 9.96 say 10 mm Ans.
...(Taking C = 0.1 and σt = 42 MPa = 42 N/mm2)
3. Size of studs for the cylinder head
Let d = Nominal diameter of the stud in mm,
dc = Core diameter of the stud in mm. It is usually taken as 0.84 d.
σt = Tensile stress for the material of the stud which is usually nickel
ns = Number of studs.
We know that the force acting on the cylinder head (or on the studs)
= × D 2 × p = (115) 2 3.15 = 32 702 N ...(i)
The number of studs (ns ) are usually taken between 0.01 D + 4 (i.e. 0.01 × 115 + 4 = 5.15) and
0.02 D + 4 (i.e. 0.02 × 115 + 4 = 6.3). Let us take ns = 6.
We know that resisting force offered by all the studs
= ns × ( dc )2 σt = 6 × (0.84d ) 2 65 = 216 d 2 N ...(ii)
...(Taking σt = 65 MPa = 65 N/mm2)
From equations (i) and (ii),
d 2 = 32 702 / 216 = 151 or d = 12.3 say 14 mm
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The pitch circle diameter of the studs (Dp) is taken D + 3d.
∴ Dp = 115 + 3 × 14 = 157 mm
We know that pitch of the studs
π × Dp π × 157
= = = 82.2 mm
We know that for a leak-proof joint, the pitch of the studs should lie between 19 d to 28.5 d ,
where d is the nominal diameter of the stud.
∴ Minimum pitch of the studs
= 19 d = 19 14 = 71.1 mm
and maximum pitch of the studs
= 28.5 d = 28.5 14 = 106.6 mm
Since the pitch of the studs obtained above (i.e. 82.2 mm) lies within 71.1 mm and 106.6 mm,
therefore, size of the stud (d ) calculated above is satisfactory.
∴ d = 14 mm Ans.
The piston is a disc which reciprocates within a cylinder. It is either moved by the fluid or it
moves the fluid which enters the cylinder. The main function of the piston of an internal combustion
engine is to receive the impulse from the expanding gas and to transmit the energy to the crankshaft
through the connecting rod. The piston must also disperse a large amount of heat from the combustion
chamber to the cylinder walls.
Fig. 32.3. Piston for I.C. engines (Trunk type).
- Internal Combustion Engine Parts n 1133
The piston of internal combustion engines are usually of trunk type as shown in Fig. 32.3. Such
pistons are open at one end and consists of the following parts :
1. Head or crown. The piston head or crown may be flat, convex or concave depending upon
the design of combustion chamber. It withstands the pressure of gas in the cylinder.
2. Piston rings. The piston rings are used to seal the cyliner in order to prevent leakage of the
gas past the piston.
3. Skirt. The skirt acts as a bearing for the side thrust of the connecting rod on the walls of
4. Piston pin. It is also called gudgeon pin or wrist pin. It is used to connect the piston to the
32.6 Design Considerations for a Piston
In designing a piston for I.C. engine, the following points should be taken into consideration :
1. It should have enormous strength to withstand the high gas pressure and inertia forces.
2. It should have minimum mass to minimise the inertia forces.
3. It should form an effective gas and oil sealing of the cylinder.
4. It should provide sufficient bearing area to prevent undue wear.
5. It should disprese the heat of combustion quickly to the cylinder walls.
6. It should have high speed reciprocation without noise.
7. It should be of sufficient rigid construction to withstand thermal and mechanical distortion.
8. It should have sufficient support for the piston pin.
32.7 Material for Pistons
The most commonly used materials for pistons of I.C. engines are cast iron, cast aluminium,
forged aluminium, cast steel and forged steel. The cast iron pistons are used for moderately rated
2. Side view
1. Front view
Twin cylinder airplane engine of 1930s.
- 1134 n A Textbook of Machine Design
engines with piston speeds below 6 m / s and aluminium alloy pistons are used for highly rated en-
gines running at higher piston sppeds. It may be noted that
1. Since the *coefficient of thermal expansion for aluminium is about 2.5 times that of cast iron,
therefore, a greater clearance must be provided between the piston and the cylinder wall (than with
cast iron piston) in order to prevent siezing of the piston when engine runs continuously under heavy
loads. But if excessive clearance is allowed, then the piston will develop ‘piston slap’ while it is cold
and this tendency increases with wear. The less clearance between the piston and the cylinder wall
will lead to siezing of piston.
2. Since the aluminium alloys used for pistons have high **heat conductivity (nearly four
times that of cast iron), therefore, these pistons ensure high rate of heat transfer and thus keeps
down the maximum temperature difference between the centre and edges of the piston head or
Notes: (a) For a cast iron piston, the temperature at the centre of the piston head (TC) is about 425°C to 450°C
under full load conditions and the temperature at the edges of the piston head (TE) is about 200°C to 225°C.
(b) For aluminium alloy pistons, TC is about 260°C to 290°C and TE is about 185°C to 215°C.
3. Since the aluminium alloys are about ***three times lighter than cast iron, therfore, its
mechanical strength is good at low tempreatures, but they lose their strength (about 50%) at temperatures
above 325°C. Sometimes, the pistons of aluminium alloys are coated with aluminium oxide by an
32.8 Piston Head or Crown
The piston head or crown is designed keeping in view the following two main considerations, i.e.
1. It should have adequate strength to withstand the straining action due to pressure of explosion
inside the engine cylinder, and
2. It should dissipate the heat of combustion to the cylinder walls as quickly as possible.
On the basis of first consideration of straining action, the thickness of the piston head is determined
by treating it as a flat circular plate of uniform thickness, fixed at the outer edges and subjected to a
uniformly distributed load due to the gas pressure over the entire cross-section.
The thickness of the piston head (tH ), according to Grashoff’s formula is given by
3 p.D 2
tH = (in mm) ...(i)
where p = Maximum gas pressure or explosion pressure in N/mm2,
D = Cylinder bore or outside diameter of the piston in mm, and
σt = Permissible bending (tensile) stress for the material of the piston in
MPa or N/mm2. It may be taken as 35 to 40 MPa for grey cast iron,
50 to 90 MPa for nickel cast iron and aluminium alloy and 60 to
100 MPa for forged steel.
On the basis of second consideration of heat transfer, the thickness of the piston head should be
such that the heat absorbed by the piston due combustion of fuel is quickly transferred to the cylinder
walls. Treating the piston head as a flat ciucular plate, its thickness is given by
tH = 12.56k (T − T ) (in mm) ...(ii)
* The coefficient of thermal expansion for aluminium is 0.24 × 10–6 m / °C and for cast iron it is 0.1 × 10–6 m / °C.
** The heat conductivity for aluminium is 174.75 W/m/°C and for cast iron it is 46.6 W/m /°C.
*** The density of aluminium is 2700 kg / m3 and for cast iron it is 7200 kg / m3.
- Internal Combustion Engine Parts n 1135
where H = Heat flowing through the piston head in kJ/s or watts,
k = Heat conductivity factor in W/m/°C. Its value is 46.6 W/m/°C for
grey cast iron, 51.25 W/m/°C for steel and 174.75 W/m/°C for
TC = Temperture at the centre of the piston head in °C, and
TE = Temperature at the edges of the piston head in °C.
The temperature difference (TC – TE) may be taken as 220°C for cast iron and 75°C for aluminium.
The heat flowing through the positon head (H) may be deternined by the following expression, i.e.,
H = C × HCV × m × B.P. (in kW)
where C = Constant representing that portion of the heat supplied to the engine
which is absorbed by the piston. Its value is usually taken
HCV = Higher calorific value of the fuel in kJ/kg. It may be taken as
45 × 103 kJ/kg for diesel and 47 × 103 kJ/ kg for petrol,
m = Mass of the fuel used in kg per brake power per second, and
B.P. = Brake power of the engine per cylinder
Notes : 1. The thickness of the piston head (tH) is calculated by using equations (i) and (ii) and larger of the two
values obtained should be adopted.
2. When tH is 6 mm or less, then no ribs are required to strengthen the piston head against gas loads. But
when tH is greater then 6 mm, then a suitable number of ribs at the centre line of the boss extending around the
skirt should be provided to distribute the side thrust from the connecting rod and thus to prevent distortion of the
skirt. The thickness of the ribs may be takes as tH / 3 to tH / 2.
3. For engines having length of stroke to cylinder bore (L / D) ratio upto 1.5, a cup is provided in the top
of the piston head with a radius equal to 0.7 D. This is done to provide a space for combustion chamber.
32.9 Piston Rings
The piston rings are used to impart the necessary radial pressure to maintain the seal between
the piston and the cylinder bore. These are usually made of grey cast iron or alloy cast iron because of
their good wearing properties and also they retain spring characteristics even at high temperatures.
The piston rings are of the following two types :
1. Compression rings or pressure rings, and
2. Oil control rings or oil scraper.
The compression rings or pressure rings are inserted in the grooves at the top portion of the
piston and may be three to seven in number. These rings also transfer heat from the piston to the
cylinder liner and absorb some part of the piston fluctuation due to the side thrust.
The oil control rings or oil scrapers are provided below the compression rings. These rings
provide proper lubrication to the liner by allowing sufficient oil to move up during upward stroke and
at the same time scraps the lubricating oil from the surface of the liner in order to minimise the flow
of the oil to the combustion chamber.
The compression rings are usually made of rectangular cross-section and the diameter of the
ring is slightly larger than the cylinder bore. A part of the ring is cut- off in order to permit it to go into
the cylinder against the liner wall. The diagonal cut or step cut ends, as shown in Fig. 32.4 (a) and (b)
respectively, may be used. The gap between the ends should be sufficiently large when the ring is put
cold so that even at the highest temperature, the ends do not touch each other when the ring expands,
otherwise there might be buckling of the ring.
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Fig. 32.4. Piston rings.
The radial thickness (t1) of the ring may be obtained by considering the radial pressure between
the cylinder wall and the ring. From bending stress consideration in the ring, the radial thickness is
t1 = D σ
where D = Cylinder bore in mm,
pw = Pressure of gas on the cylinder wall in N/mm2. Its value is limited
from 0.025 N/mm2 to 0.042 N/mm2, and
σt = Allowable bending (tensile) stress in MPa. Its value may be taken
from 85 MPa to 110 MPa for cast iron rings.
The axial thickness (t2 ) of the rings may be taken as 0.7 t1 to t1.
The minimum axial thickness (t2 ) may also be obtained from the following empirical relation:
t2 = 10 n
where nR = Number of rings.
The width of the top land (i.e. the distance from the top of the piston to the first ring groove) is
made larger than other ring lands to protect the top ring from high temperature conditions existing at
the top of the piston,
∴ Width of top land,
b1 = tH to 1.2 tH
The width of other ring lands (i.e. the distance between the ring grooves) in the piston may be
made equal to or slightly less than the axial thickness of the ring (t2).
∴ Width of other ring lands,
b2 = 0.75 t2 to t2
The depth of the ring grooves should be more than the depth of the ring so that the ring does not
take any piston side thrust.
The gap between the free ends of the ring is given by 3.5 t1 to 4 t1. The gap, when the ring is in
the cylinder, should be 0.002 D to 0.004 D.
32.10 Piston Barrel
It is a cylindrical portion of the piston. The maximum thickness (t3) of the piston barrel may be
obtained from the following empirical relation :
t3 = 0.03 D + b + 4.5 mm
- Internal Combustion Engine Parts n 1137
where b = Radial depth of piston ring groove which is taken as 0.4 mm larger
than the radial thickness of the piston ring (t1)
= t1 + 0.4 mm
Thus, the above relation may be written as
t3 = 0.03 D + t1 + 4.9 mm
The piston wall thickness (t4) towards the open end is decreased and should be taken as 0.25 t3
to 0.35 t3.
32.11 Piston Skirt
The portion of the piston below the ring section is known as piston skirt. In acts as a bearing for
the side thrust of the connecting rod. The length of the piston skirt should be such that the bearing
pressure on the piston barrel due to the side thrust does not exceed 0.25 N.mm2 of the projected area
for low speed engines and 0.5 N/mm2 for high speed engines. It may be noted that the maximum
thrust will be during the expansion stroke. The side thrust (R) on the cylinder liner is usually taken as
1/10 of the maximum gas load on the piston.
Valve Cylinder head
Magneto Ignition leads
Spur gear on end of
1000 cc twin -cylinder motorcycle engine.
We know that maximum gas load on the piston,
P = p×
∴ Maximum side thrust on the cylinder,
R = P/10 = 0.1 p × ...(i)
where p = Maximum gas pressure in N/mm2, and
D = Cylinder bore in mm.
The side thrust (R) is also given by
R = Bearing pressure × Projected bearing area of the piston skirt
= pb × D × l
where l = Length of the piston skirt in mm. ...(ii)
- 1138 n A Textbook of Machine Design
From equations (i) and (ii), the length of the piston skirt (l) is determined. In actual practice, the
length of the piston skirt is taken as 0.65 to 0.8 times the cylinder bore. Now the total length of the
piston (L) is given by
L = Length of skirt + Length of ring section + Top land
The length of the piston usually varies between D and 1.5 D. It may be noted that a longer piston
provides better bearing surface for quiet running of the engine, but it should not be made unnecessarily
long as it will increase its own mass and thus the inertia
32.12 Piston Pin
The piston pin (also called gudgeon pin or wrist pin)
is used to connect the piston and the connecting rod. It is
usually made hollow and tapered on the inside, the smallest Fig.32.5. Piston pin.
inside diameter being at the centre of the pin, as shown in Fig. 32.5. The piston pin passes through the
bosses provided on the inside of the piston skirt and the bush of the small end of the connecting rod.
The centre of piston pin should be 0.02 D to 0.04 D above the centre of the skirt, in order to off-set
the turning effect of the friction and to obtain uniform distribution of pressure between the piston and
the cylinder liner.
The material used for the piston pin is usually case hardened steel alloy containing nickel,
chromium, molybdenum or vanadium having tensile strength from 710 MPa to 910 MPa.
Fig. 32.6. Full floating type piston pin.
The connection between the piston pin and the small end of the connecting rod may be made
either full floating type or semi-floating type. In the full floating type, the piston pin is free to turn
both in the *piston bosses and the bush of the small end of the connecting rod. The end movements of
the piston pin should be secured by means of spring circlips, as shown in Fig. 32.6, in order to prevent
the pin from touching and scoring the cylinder liner.
In the semi-floating type, the piston pin is either free to turn in the piston bosses and rigidly
secured to the small end of the connecting rod, or it is free to turn in the bush of the small end of
the connecting rod and is rigidly secured in the piston bosses by means of a screw, as shown in
The piston pin should be designed for the maximum gas load or the inertia force of the piston,
whichever is larger. The bearing area of the piston pin should be about equally divided between the
piston pin bosses and the connecting rod bushing. Thus, the length of the pin in the connecting rod
bushing will be about 0.45 of the cylinder bore or piston diameter (D), allowing for the end clearance
* The mean diameter of the piston bosses is made 1.4 d0 for cast iron pistons and 1.5 d0 for aluminium
pistons, where d0 is the outside diameter of the piston pin. The piston bosses are usually tapered, increasing
the diameter towards the piston wall.
- Internal Combustion Engine Parts n 1139
of the pin etc. The outside diameter of the piston pin (d0 ) is determined by equating the load on the
piston due to gas pressure (p) and the load on the piston pin due to bearing pressure ( pb1 ) at the small
end of the connecting rod bushing.
(a) Piston pin secured to the small (b) Piston pin secured to the boss
end of the connecting rod. of the piston.
Fig. 32.7. Semi-floating type piston pin.
Let d0 = Outside diameter of the piston pin in mm
l1 = Length of the piston pin in the bush of the small end of the connecting
rod in mm. Its value is usually taken as 0.45 D.
pb1 = Bearing pressure at the small end of the connecting rod bushing in
N/mm2. Its value for the bronze bushing may be taken as 25 N/mm2.
We know that load on the piston due to gas pressure or gas load
= × p ..(i)
and load on the piston pin due to bearing pressure or bearing load
= Bearing pressure × Bearing area = pb1 × d0 × l1, ...(ii)
From equations (i) and (ii), the outside diameter of the piston pin (d0) may be obtained.
The piston pin may be checked in bending by assuming the gas load to be uniformly distributed
over the length l1 with supports at the centre of
the bosses at the two ends. From Fig. 32.8, we
find that the length between the supports,
D − l1 l + D
l2 = l1 + = 1
Now maximum bending moment at the
centre of the pin,
P l2 P l1 l1
M= 2 × 2 − l × 2 × 4
P l2 P l1
= × − ×
2 2 2 4
P l1 + D P l1 Fig. 32.8
=2 − ×
2×2 2 4
P. l1 P.D P. l1 P.D
= + − = ...(iii)
8 8 8 8
- 1140 n A Textbook of Machine Design
We have already discussed that the piston pin is made hollow. Let d0 and di be the outside and
inside diameters of the piston pin. We know that the section modulus,
π ( d 0 ) 4 − ( di ) 4
Z = 32
We know that maximum bending moment,
π ( d 0 ) 4 − ( di ) 4
M = Z × σb = 32 σb
where σb = Allowable bending stress for the material of the piston pin. It is
usually taken as 84 MPa for case hardened carbon steel and
140 MPa for heat treated alloy steel.
Assuming di = 0.6 d0, the induced bending stress in the piston pin may be checked.
Spring Cam chain is driven by
pushed round by
Waste gases out
Fuel flows in when Crankshaft
Another view of a single cylinder 4-stroke petrol engine.
Example 32.2. Design a cast iron piston for a single acting four stroke engine for the following
Cylinder bore = 100 mm ; Stroke = 125 mm ; Maximum gas pressure = 5 N/mm2 ; Indicated
mean effective pressure = 0.75 N/mm2 ; Mechanical efficiency = 80% ; Fuel consumption = 0.15 kg
per brake power per hour ; Higher calorific value of fuel = 42 × 103 kJ/kg ; Speed = 2000 r.p.m.
Any other data required for the design may be assumed.
Solution. Given : D = 100 mm ; L = 125 mm = 0.125 m ; p = 5 N/mm2 ; pm = 0.75 N/mm2;
ηm = 80% = 0.8 ; m = 0.15 kg / BP / h = 41.7 × 10–6 kg / BP / s; HCV = 42 × 103 kJ / kg ;
N = 2000 r.p.m.
The dimensions for various components of the piston are determined as follows :
1. Piston head or crown
The thickness of the piston head or crown is determined on the basis of strength as well as on the
basis of heat dissipation and the larger of the two values is adopted.
- Internal Combustion Engine Parts n 1141
We know that the thickness of piston head on the basis of strength,
3 p . D2 3 × 5 (100) 2
tH = = = 15.7 say 16 mm
16 σt 16 × 38
...(Taking σt for cast iron = 38 MPa = 38 N/mm2)
Since the engine is a four stroke engine, therefore, the number of working strokes per minute,
n = N / 2 = 2000 / 2 = 1000
and cross-sectional area of the cylinder,
π D2 π (100) 2
A = = = 7855 m m 2
We know that indicated power,
p m . L. A.n 0.75 × 0.125 × 7855 × 1000
IP = = = 12 270 W
= 12.27 kW
∴ Brake power, BP = IP × ηm = 12.27 × 0.8 = 9.8 kW ...(∴ ηm = BP / IP)
We know that the heat flowing through the piston head,
H = C × HCV × m × BP
= 0.05 × 42 × 103 × 41.7 × 10–6 × 9.8 = 0.86 kW = 860 W
....(Taking C = 0.05)
∴Thickness of the piston head on the basis of heat dissipation,
tH = 12.56 k (T − T ) = 12.56 × 46.6 × 220 = 0.0067 m = 6.7 mm
...(∵ For cast iron , k = 46.6 W/m/°C, and TC – TE = 220°C)
Taking the larger of the two values, we shall adopt
tH = 16 mm Ans.
Since the ratio of L / D is 1.25, therefore a cup in the top of the piston head with a radius equal
to 0.7 D (i.e. 70 mm) is provided.
2. Radial ribs
The radial ribs may be four in number. The thickness of the ribs varies from tH / 3 to tH / 2.
∴ Thickness of the ribs, tR = 16 / 3 to 16 / 2 = 5.33 to 8 mm
Let us adopt tR = 7 mm Ans.
3. Piston rings
Let us assume that there are total four rings (i.e. nr = 4) out of which three are compression rings
and one is an oil ring.
We know that the radial thickness of the piston rings,
3 pw 3 × 0.035
t1 = D = 100 = 3.4 mm
...(Taking pw = 0.035 N/mm2, and σt = 90 MPa)
and axial thickness of the piston rings
t2 = 0.7 t1 to t1 = 0.7 × 3.4 to 3.4 mm = 2.38 to 3.4 mm
Let us adopt t2 = 3 mm
- 1142 n A Textbook of Machine Design
We also know that the minimum axial thickness of the pistion ring,
t2 = = = 2.5 m m
10 n r 10 × 4
Thus the axial thickness of the piston ring as already calculated (i.e. t2 = 3 mm)is satisfactory. Ans.
The distance from the top of the piston to the first ring groove, i.e. the width of the top land,
b1 = tH to 1.2 tH = 16 to 1.2 × 16 mm = 16 to 19.2 mm
and width of other ring lands,
b2 = 0.75 t2 to t2 = 0.75 × 3 to 3 mm = 2.25 to 3 mm
Let us adopt b1 = 18 mm ; and b2 = 2.5 mm Ans.
We know that the gap between the free ends of the ring,
G1 = 3.5 t1 to 4 t1 = 3.5 × 3.4 to 4 × 3.4 mm = 11.9 to 13.6 mm
and the gap when the ring is in the cylinder,
G2 = 0.002 D to 0.004 D = 0.002 × 100 to 0.004 × 100 mm
= 0.2 to 0.4 mm
Let us adopt G1 = 12.8 mm ; and G2 = 0.3 mm Ans.
4. Piston barrel
Since the radial depth of the piston ring grooves (b) is about 0.4 mm more than the radial
thickness of the piston rings (t1), therefore,
b = t1 + 0.4 = 3.4 + 0.4 = 3.8 mm
We know that the maximum thickness of barrel,
t3 = 0.03 D + b + 4.5 mm = 0.03 × 100 + 3.8 + 4.5 = 11.3 mm
and piston wall thickness towards the open end,
t4 = 0.25 t3 to 0.35 t3 = 0.25 × 11.3 to 0.35 × 11.3 = 2.8 to 3.9 mm
Let us adopt t4 = 3.4 mm
5. Piston skirt
Let l = Length of the skirt in mm.
We know that the maximum side thrust on the cylinder due to gas pressure ( p ),
π D2 π (100) 2
R = µ× × p = 0.1 × × 5 = 3928 N
...(Taking µ = 0.1)
We also know that the side thrust due to bearing pressure on the piston barrel ( pb ),
R = pb × D × l = 0.45 × 100 × l = 45 l N
...(Taking pb = 0.45 N/mm2)
From above, we find that
45 l = 3928 or l = 3928 / 45 = 87.3 say 90 mm Ans.
∴ Total length of the piston ,
L = Length of the skirt + Length of the ring section + Top land
= l + (4 t2 + 3b2) + b1
= 90 + (4 × 3 + 3 × 3) + 18 = 129 say 130 mm Ans.
6. Piston pin
Let d0 = Outside diameter of the pin in mm,
l1 = Length of pin in the bush of the small end of the connecting rod in
- Internal Combustion Engine Parts n 1143
pb1 = Bearing pressure at the small end of the connecting rod bushing in
N/mm2. It value for bronze bushing is taken as 25 N/mm2.
We know that load on the pin due to bearing pressure
= Bearing pressure × Bearing area = pb1 × d0 × l1
= 25 × d0 × 0.45 × 100 = 1125 d0 N ...(Taking l1 = 0.45 D)
We also know that maximum load on the piston due to gas pressure or maximum gas load
π D2 π (100) 2
= × p= × 5 = 39 275 N
From above, we find that
1125 d0 = 39 275 or d0 = 39 275 / 1125 = 34.9 say 35 mm Ans.
The inside diameter of the pin (di) is usually taken as 0.6 d0.
∴ di = 0.6 × 35 = 21 mm Ans.
Let the piston pin be made of heat treated alloy steel for which the bending stress ( σ b )may be
taken as 140 MPa. Now let us check the induced bending stress in the pin.
We know that maximum bending moment at the centre of the pin,
P .D 39 275 × 100
M = = = 491 × 10 3 N-mm
We also know that maximum bending moment (M),
π (d0 )4 − (di )4 π (35)4 − (21) 4
σb = σb = 3664 σb
491 × 103 =
∴ σ b = 491 × 10 3 / 3664 = 134 N/mm2 or MPa
Since the induced bending stress in the pin is less than the permissible value of 140 MPa (i.e.
140 N/mm2), therefore, the dimensions for the pin as calculated above (i.e. d0 = 35 mm and di = 21 mm)
Fan blows air over
engine to cool it
Air filter stops dust
and dirt from being
sucked into engine
Disk brake Spark plug Twin rotors
German engineer Fleix Wankel (1902-88) built a rotary engine in 1957. A triangular piston turns inside a
chamber through the combustion cycle.
- 1144 n A Textbook of Machine Design
32.13 Connecting Rod
The connecting rod is the intermediate member between the piston and the crankshaft. Its primary
function is to transmit the push and pull from the piston pin to the crankpin and thus convert the
reciprocating motion of the piston into the rotary motion of the crank. The usual form of the connecting
rod in internal combustion engines is shown in Fig. 32.9. It consists of a long shank, a small end and a
big end. The cross-section of the shank may be rectangular, circular, tubular, I-section or H-section.
Generally circular section is used for low speed engines while I-section is preferred for high speed engines.
Fig. 32.9. Connecting rod.
The *length of the connecting rod ( l ) depends upon the ratio of l / r, where r is the radius of
crank. It may be noted that the smaller length will decrease the ratio l / r. This increases the angularity
of the connecting rod which increases the side thrust of the piston against the cylinder liner which in
turn increases the wear of the liner. The larger length of the connecting rod will increase the ratio
l / r. This decreases the angularity of the connecting rod and thus decreases the side thrust and the
resulting wear of the cylinder. But the larger length of the connecting rod increases the overall height
of the engine. Hence, a compromise is made and the ratio l / r is generally kept as 4 to 5.
The small end of the connecting rod is usually made in the form of an eye and is provided with
a bush of phosphor bronze. It is connected to the piston by means of a piston pin.
The big end of the connecting rod is usually made split (in two **halves) so that it can be
mounted easily on the crankpin bearing shells. The split cap is fastened to the big end with two cap
bolts. The bearing shells of the big end are made of steel, brass or bronze with a thin lining (about
0.75 mm) of white metal or babbit metal. The wear of the big end bearing is allowed for by inserting
thin metallic strips (known as shims) about 0.04 mm thick between the cap and the fixed half of the
connecting rod. As the wear takes place, one or more strips are removed and the bearing is trued up.
* It is the distance between the centres of small end and big end of the connecting rod.
** One half is fixed with the connecting rod and the other half (known as cap) is fastened with two cap bolts.